Stepper motor driven proportional actuator

ABSTRACT

A low energy stepper motor driven actuator that eliminates the need for a position sensor is provided. The stepper motor rotates a cam in a control piston valve. The cam rotation increases the gap between the cam and nozzle on one side of the cam and decreases the gap between the cam and nozzle on the other side. The gap differences affect the pressures on the control piston ends, which forces the piston in the direction that will re-equalize the cam-nozzle gaps. As a result, piston moves to a position such that the head or rod of the actuator piston receives high pressure flow, thereby moving the actuator. Movement of the actuator rod provides mechanical feedback to the cam, causing the cam to move back to its mechanical null position.

FIELD OF THE INVENTION

This invention pertains to actuators, and more particularly tosensorless actuators.

BACKGROUND OF THE INVENTION

Conventional actuator systems employ a closed loop position controlsystem. These systems include a position sensor for actuator positionfeedback and either an integrating controller or proportional controllerused for control. The integrating controller assures that the steadystate sensed position matches the commanded position. However, theactual position versus commanded position is still susceptible toinaccuracies of the position sensor gain and position (i.e., calibrationof the position sensor to the valve position), the position sensordemodulator accuracy, channel-channel tracking and digital resolution.The proportional controller is susceptible to the above inaccuracies aswell as an allowed steady state error that is a function of disturbancemagnitudes and the proportional gain of the controller.

Regardless of the controller type, the accuracy of the system is veryhighly dependent on the position sensor accuracy. For precise meteringapplications such as in aircraft systems, the position sensors need tobe very accurate and have high resolution. While very accurate, theposition sensors are typically very expensive, both in terms of time andcost. They are relatively difficult to interface with due to themechanical interface, the hydraulic interface, the number of small gaugewires, complicated demodulation circuitry, etc. Position sensors arealso prone to failure due to the reliability of small gauge wires. Thisfailure mode leads to dual channel requirements (i.e., two separateposition sensors, drivers, and motor control) and additional cost inorder to meet reliability requirements.

Elimination of the position feedback sensor will save money and weight.However, the lack of position feedback and the closed loop controllermeans that the effects of disturbances and/or the variations in forwardpath gain that are sensed and/or compensated in the closed loopcontroller will no longer be sensed and/or compensated. To negate theseadverse effects, the magnitude of the disturbances should be minimized,the inherent disturbance rejection characteristics of the forward pathshould be maximized and the gain accuracy of the forward path should bemade insensitive to the environment. In other words, the forward pathmust be “robust.” The forward path must also be strictly proportionalsince there is no feedback to prevent the divergence that would occurwith an integrating forward path.

Open loop, proportional electro-hydraulic servo valve (EHSV) basedactuator systems use a low energy torque motor that controls hydraulicsthat drive the actuator. The motor used has high speed but very lowtorque. The low torque levels result in the motor (and thus theactuator) being substantially affected by relatively small DC torquedisturbances. For example, isolation seals, relaxation of torsion springpreload, magnet MMF (magnetomotive force) variations, variations in fluxpath reluctance, discrete steps in nozzle pressure feedback forces,thermal induced movement of parts, etc. can affect the torque motor. Therelatively undamped torque motor also does not support good dynamictorque disturbance rejection (e.g., current transient, vibration, etc.)and creates resonance issues. The actuator position is fed back to themotor via springs. This indirect position feedback technique does notprovide adequate load disturbance rejection for most applications.

What is needed is a system that overcomes the problems of sensorlessactuators as discussed above. The invention provides a system with suchfeatures. These and other advantages of the invention, as well asadditional inventive features, will be apparent from the description ofthe invention provided herein.

BRIEF SUMMARY OF THE INVENTION

The invention provides a stepper motor driven proportional actuator thateliminates the need for a position feedback sensor. The stepper motor isused to drive a cam that is designed such that the cross-cam distance onthe nozzle-nozzle centerline of the fuel metering valve is a constantfor any operational cam angle. Additionally, the tangent to the camsurface is perpendicular to the nozzle-nozzle centerline, therebyallowing the cam to contact and push on the nozzles if needed.

The stepper motor drives a gearbox connected to the cam. The gearbox inone embodiment is a planetary gear system having an outer ring gearconnected in mesh relationship to sun gears, which are connectedtogether with a sun frame and are in mesh relationship with a piniongear. The outer ring gear is rotatably connected to a rack that isintegrated to the actuator piston. The pinion gear is integral to thestepper motor rotor. When the stepper motor turns, the resultant piniongear rotation cause the sun gears and sun frame to rotate. An outputshaft that is attached to the sun frame and cam rotates with it. The camrotation opens (or closes) cam/nozzle gaps causing the control piston totranslate, thereby opening the head and rod ports to supply or draincausing the actuator piston and rack to stroke. The rack provides directactuator position feedback to the outer ring gear and causes the outerring gear to rotate. The rotation of the ring gear causes the sun gearsand the sun frame to rotate back to their original position, therebyrotating the cam and translating the control valve to its mechanicalnull position.

Other aspects, objectives and advantages of the invention will becomemore apparent from the following detailed description when taken inconjunction with the accompanying drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a cross-sectional view of the actuator system in accordancewith the teachings of the present invention;

FIG. 2 is a partial cross-sectional view of the actuator system of FIG.1 with the stepper motor shown as a separate component for clarity andthe valve piston at a centered position;

FIG. 3 is a schematic view of the actuator system of FIG. 2 illustratingthe cam-rack interaction;

FIG. 4 is a partial cross-sectional view of the actuator system of FIG.2 with the piston at a position such that flow drives the actuator inthe retract direction with the against the retract stop; and

FIG. 5 is a partial cross-sectional view of the actuator system of FIG.2 with the piston at a position such that flow drives the actuator inthe extend direction with the against the extend stop.

While the invention will be described in connection with certainpreferred embodiments, there is no intent to limit it to thoseembodiments. On the contrary, the intent cover all alternatives,modifications and equivalents as included within the spirit and theinvention as defined by the appended claims.

DETAILED DESCRIPTION OF THE INVENTION

The invention provides a stepper motor driven actuator system thateliminates the need for a position sensor and position feedback. Thehydraulic amplification that is typically provided by an EHSV flappervalve is eliminated and replaced with a constant gain cam-nozzleamplification-tracking system. The combination of the cam-nozzle,stepper motor, and a gearbox in communication with the rack of theactuator piston provides an accurate and robust actuation positioningsystem. One feature of the invention is that it can provide a“fail-fixed” system, that is, a system that maintains it's lastcommanded position in the event of electrical power failure.

Turning to the drawings, wherein like reference numerals refer to likeelements, the invention provides a stepper motor driven robustproportional actuator. With reference to FIGS. 1 to 3, a stepper motor100 is used to drive cam 102. The stepper motor drives a planetary gearsystem 104 where the ring gear 106 is in mesh relation to rack 108. Thepinion gear 110 is integral to the stepper motor rotor 112. When thestepper motor 100 is rotated, the pinion gear 110 rotates. The piniongear 110 rotation causes the sun gears 114 and sun frame 116 to rotate.The output shaft 118 is attached to the sun frame 116 and rotates withit. Similarly, the cam 102 that is attached to the output shaft 118rotates with the output shaft 118.

The cam rotation increases the gap between the cam 102 and nozzle 120 onone side of the cam 102 and decreases the gap between the cam 102 andnozzle 120 on the other side. The differences in the gaps affect the Pz1and Pz2 pressures on the ends 124 of the control piston 122 so as toforce the control piston 122 in the direction that will re-equalize thecam-nozzle gaps. The control piston translation opens the head port 126and rod port 128 to supply or drain, thereby causing the actuator piston130 and rack 108 to stroke. The rack 108 provides direct actuatorposition feedback to the ring gear 106, causing the ring gear 106 torotate. The ring gear rotation causes the sun gears 114 and sun frame116 to rotate back to their original position, thereby rotating the camand translating the control piston 122 to the mechanical null position(i.e., the center position).

When the cam 102 is in the center position, the hydraulic flow willenter port 134, pass through the cam-nozzle-orifice system (i.e., aroundcam 102 and through nozzles 120 and corresponding orifices), enter line136, and then drain out through Pb port 138 due to the lower pressure inthe Pb drain. It should be noted that the direction of flow is from line134 and into the nozzles 120 via the cam-nozzle gap (i.e., “flow in”) ascompared to conventional valves where flow is from the piston ends 124out of the nozzle 120 (i.e., “flow out”).

Note that when the cam 102 is positioned such that the control piston122 is towards the left-most position 140 in the control valve body 132,the supply port 134 is opened to the head port 126 (see FIG. 5). Whenthis occurs, the hydraulic flow passes through port 134, out head port126 and returns through rod port 128 and discharges out Pb port 138.When the cam 102 is positioned such that the control piston 122 istowards the right-most position 142 in the control valve body 132, thesupply port 134 is opened to the rod port 128 (see FIG. 4). Thehydraulic flow passes through port 134, out rod port 128 and returnsthrough head port 126 and discharges out Pb port 144.

During normal operation with a properly sized hydraulic andelectromechanical system, it is unlikely that the control piston 122will be at either its left-most position 140 or its right-most position142 (as respectively shown in FIGS. 5 and 4) due to the response of thesystem. In FIG. 4, the hydraulics are driving the actuator in theretract direction but it is against the retract stop. FIG. 5 depicts thehydraulics driving the actuator in the extend direction but the actuatoris against the extend stop. Generally, as the stepper motor 100 rotatesthe cam 102, the control piston 122 begins to move and flow enters oneof the head port 126 or rod port 128. As the control piston 122continues to move as the cam 102 is rotated, the port through which flowenters (i.e., head port 126 or rod port 128) opens wider, thusincreasing flow. As the flow pushes actuator piston 130, rod 108 moves,thus rotating ring gear 106 as described above. The rotation of ringgear 106 by rod 108 results in the cam 102 and control piston 122translating to the mechanical null position, thus preventing furtherflow to the actuator. The result is a proportional tracking of theactuator 130 to the motor rotor 112. As long as the dynamics of thesystem are sufficient fast so as to keep up with the input from themotor 100, the actuator 130 will track the motor 100 commands withrelatively small transient rotations of shaft 11 8, cam 102 andtranslations of control piston 122.

The primary disturbance to the system is the force input to theactuator. Any movement of the actuator piston 130 will cause rack 108translation and ring gear 106 rotation. Any ring gear movement resultsin cam 102 rotation due to the precision gearbox system 104. The highpressure gain of the system assures control piston 122 movement for anycam 102 rotation. The high pressure gain of the control valve ports 126& 128 coupled with the large head/rod areas will result in the requiredresistive force with minimal position error.

An example of the stiffness of the system is provided below. Assume forpurposes of discussion that there are 103 gear teeth on the innerdiameter of the ring gear 106 (approximately 130 teeth on the outerdiameter), 45 gear teeth oil each sun gear 114, 13 gear teeth on thepinion gear 110, 33 teeth per inch on rack 108 and the motor pinionrotating approximately 1.73 degrees per step of the stepper motor 100.With the proper cam sizing as described in U.S. application Ser. No.11/094,099, filed Apr. 27, 2005, hereby incorporated by reference in itsentirety, there is 418 steps per inch of actuator stroke. A 0.001 inchactuator stroke error is equivalent to 0.091 degrees of ring rotation.0.091 degrees of ring rotation leads to 0.081 degrees of cam rotation(neglecting gearbox slop). The cam rise of 0.004 in/degree yields0.000324 inches of control valve stroke. Assuming a control portpressure gain of approximately 125000 psid/in, the actuator stroke errorequates to a dP of 40 psid. Assuming an actuator with a 5 in² head areaand a 4.25 in² rod area, Pb=100 and Ps=600 psid, the resulting change inforce is approximately 185 lbs. The resulting disturbance rejection isapproximately 185,000 lbs/in (neglecting any rack and gearbox backlash).A 500 lb external force would move the actuator 0.0027 inches or 0.09%stk. As can be seen from the foregoing, the actuator system of theinvention is very stiff.

The stepper motor system is a relatively low energy motor coupled torelatively high energy hydraulics. The stepper motor in combination witha gearbox provides the capability to decrease stepper motor speed andincrease its torque while staying at the same energy level. This isaccomplished by proper selection of the motor stator and rotor toothcount and the gearbox ratio. This can be used to increase the motortorque, decrease it's susceptibility to torque disturbances and stillkeep the motor fast enough to handle dynamic operation. The steppermotor has nearly perfect gain and is essentially unaffected by torquedisturbances due to higher torque capability, the gear box torqueamplification and the inherent detent feature of the stepper motor. Theround, symmetrical, balanced construction of the stepper motor is inessence unaffected by vibration and temperature variations.

The precision machined placement of stator and rotor teeth provide theinherent baseline position and gain accuracy of the stepper motor. If anaccurate calibration is made, and the effects of disturbances arenegated, the need for a sensor is eliminated. This accuracy does notchange with life, is essentially constant from unit to unit, and is nota function of any calibration procedure. The round, symmetricconstruction of the stepper motor maintains this accuracy in thepresence of environment variations (e.g., temperature). Torquedisturbances at the output shaft such as dynamic seal friction, nozzlehydraulic loads, unbalanced cam mass, etc. are minimal and areessentially rejected by the precision gearbox 140 (comprising piniongear 110, ring gear 106, sun gears 114 and sun frame 116) and the highdetent torque of the motor. The detent torque prevents disturbances fromhaving any appreciable effect until they reach such a magnitude thatthey completely overpower the stepper motor. The stepper motor rotorrides on precision ball bearings and is inherently precision-balancedabout its rotation axis in the presence of translational accelerations(i.e., vibration), so the torque disturbances at the motor rotor arenegligible. The stepper motor 100 has no channel-channel tracking errordue to the fact that both channels share the same rotor-stator-pole fluxcircuit. Power transients have no effect on steady state operation andthe precision gearbox has minimal backlash. In one embodiment, thebacklash of the gearbox 140 is approximately two step increments of thestepper motor 100.

As can be seen from the foregoing, a robust stepper motor drivenproportional actuator has been described. Robustness, as used herein,refers to the ability of a system to remain accurate in the presence ofdisturbance inputs and environment variations. Disturbances lead to ashift in the entire step versus position plot and gain variations leadto changes in the slope of the plot. Disturbances are due to undesiredtorques and forces as well as imperfect calibration. Gain variations aredue to the change in output/input characteristics due to component lifeand environment. Robustness is obtained by the invention by minimizingthe magnitude of disturbances where possible, by isolating the devicefrom disturbances where necessary, maximizing the disturbance rejectioncharacteristics of the device, designing a device with minimal wearand/or whose performance is unaffected by wear, precision calibration,and inherent gain accuracy in the presence of environment variations(e.g., temperature, stray flux, vibration, pressures, etc.).

The use of the terms “a” and “an” and “the” and similar referents in thecontext of describing the invention (especially in the context of thefollowing claims) is to be construed to cover both the singular and theplural, unless otherwise indicated herein or clearly contradicted bycontext. The terms “comprising,” “having,” “including,” and “containing”are to be construed as open-ended terms (i.e., meaning “including, butnot limited to,”) unless otherwise noted. All methods described hereincan be performed in any suitable order unless otherwise indicated hereinor otherwise clearly contradicted by context. The use of any and allexamples, or exemplary language (e.g., “such as”) provided herein, isintended merely to better illuminate the invention and does not pose alimitation on the scope of the invention unless otherwise claimed. Nolanguage in the specification should be construed as indicating anynon-claimed element as essential to the practice of the invention.

Preferred embodiments of this invention are described herein, includingthe best mode known to the inventors for carrying out the invention.Variations of those preferred embodiments may become apparent to thoseof ordinary skill in the art upon reading the foregoing description. Forexample, a ball screw can be used where the ball portion is mounted inthe actuator with the screw integral to the ring gear. Actuatortranslation would cause the ring gear to rotate as described above. Theinventors expect skilled artisans to employ such variations asappropriate, and the inventors intend for the invention to be practicedotherwise than as specifically described herein. Accordingly, thisinvention includes all modifications and equivalents of the subjectmatter recited in the claims appended hereto as permitted by applicablelaw. Moreover, any combination of the above-described elements in allpossible variations thereof is encompassed by the invention unlessotherwise indicated herein or otherwise clearly contradicted by context.

1. A stepper motor driven actuator comprising: a stepper motor; a camoperatively connected to the stepper motor, the cam rotating position inresponse to stepping of the stepper motor, the cam positioned within avalve body having an inlet port, a rod port, a head port and at leastone drain port; a control piston positioned within the valve body andhaving a plurality of nozzles positioned on opposite sides of the cam inclose proximity to a surface of the cam and movable between a nullposition and flow positions, each of the plurality of nozzles having aflow path, the control piston moving as a result of a pressure imbalanceat the ends of the control piston occurring in response to a change inposition of the cam such that a difference between a first gap betweenthe surface of the cam and one of the plurality of nozzles and a secondgap between the surface of the cam and an other of the plurality ofnozzles is minimized; and an actuator having a piston integral to a rackthat is operably coupled to the cam, the piston having a first side anda second side, the first side being in fluid communication with the headport and the second side being in fluid communication with the rod port,the cam moving in response to the rack moving.
 2. The actuator of claim1 wherein the cam is connected to an output shaft, the actuator furthercomprising a gearbox connected between a rotor of the stepper motor andthe output shaft.
 3. The actuator of claim 2 wherein the gearboxcomprises: a sun frame attached to the output shaft and a plurality ofsun gears; a ring gear operatively connected to the plurality of sungears and the rack; and a pinion gear connected to the rotor andoperatively connected to the plurality of sun gears.
 4. The actuator ofclaim 3 wherein the rack provides direct actuator feedback to the ringgear.
 5. The actuator of claim 3 wherein translation of the controlpiston opens one of the head port and rod port to supply and the otherof the head port and rod to drain, thereby causing the actuator pistonand rack to stroke.
 6. The actuator of claim 3 wherein translation ofthe control piston towards a first end of the valve body opens the headport to supply and the rod port to drain.
 7. The actuator of claim 6wherein translation of the control piston towards a second end of thevalve body opens the head port to drain and the rod port to supply. 8.The actuator of claim 3 wherein the ring gear rotates in response to therack stroking, the ring gear rotation causing the plurality of sun gearsand the sun frame to rotate back to their position prior to thetranslation of the control piston, thereby rotating the cam andtranslating the control piston to a mechanical null position.
 9. Theactuator of claim 1 wherein the stepper motor is wound such that theactuator may be operated in a dual channel mode of operation.
 10. Theactuator of claim 1 wherein the control piston, the valve body and thecam provide four way operation.
 11. A stepper motor driven actuatorsystem comprising: a stepper motor; a gearbox operatively connected tothe stepper motor; a cam operatively connected to the gearbox, the camrotating position in response to stepping of the stepper motor, the campositioned within a valve body having an inlet port, a rod port, a headport and at least one drain port; a control piston positioned within thevalve body and having a pair of nozzles positioned on opposite sides ofthe cam in close proximity to a surface of the cam so as to each definea gap therebetween, each of the pair of nozzles having a flow path, thecontrol piston moving as a result of a pressure imbalance at the ends ofthe control piston occurring in response to a change in the gap betweenthe surface of the cam and each of the nozzles due to a change inposition of the cam; and an actuator having a piston integral to a rack,the piston having a first side and a second side, the first side influid communication with the head port and the second side in fluidcommunication with the rod port, the rack in mesh relationship to thegearbox, the cam moving in response to the rack moving.
 12. The actuatorsystem of claim 11 further comprising an output shaft operativelyconnected to the cam and the gearbox.
 13. The actuator system of claim12 wherein the gearbox comprises: a sun frame attached to the outputshaft and a plurality of sun gears; a ring gear operatively connected tothe plurality of sun gears and the rack; and a pinion gear connected toa rotor of the stepper motor and operatively connected to the pluralityof sun gears.
 14. The actuator system of claim 13 wherein the rackprovides direct actuator feedback to the ring gear.
 15. The actuatorsystem of claim 13 wherein the ring gear rotates in response to the rackstroking, the ring gear rotation causing the plurality of sun gears andthe sun frame to rotate back to their position prior to the translationof the control piston, thereby rotating the cam and translating thecontrol piston to a mechanical null position.
 16. The actuator system ofclaim 13 wherein rotation of the pinion gear is countered by translationof the rack.
 17. The actuator system of claim 11 wherein translation ofthe control piston towards a first end of the valve body opens the headport to supply and the rod port to drain.
 18. The actuator system ofclaim 17 wherein translation of the control piston towards a second endof the valve body opens the head port to drain and the rod port tosupply.
 19. The actuator system of claim 11 wherein each of theplurality of nozzles has a first end by the cam and a second endopposite the cam, and wherein flow enters the plurality of nozzles inthe first end and out the second end.
 20. The actuator system of claim11 where the actuator system has a disturbance rejection of at least180,000 pounds per inch.